Apparatus, method and software for use with an air conditioning cycle

ABSTRACT

A turbine for generating power has a rotor chamber, a rotor rotatable about a central axis within the rotor chamber, and at least one nozzle for supplying a fluid from a fluid supply to the rotor to thereby drive the rotor and generate power. The flow of the fluid from the nozzle exist is periodically interrupted by at least one flow interrupter means, thereby raising a pressure of the fluid inside the nozzle. A thermo-dynamic cycle is also disclosed including a compressor, a first turbine downstream of the compressor, a heat exchanger located downstream of the first turbine and operable to reject heat from the cycle to another thermodynamic cycle, an evaporator downstream of the heat exchanger and a second turbine downstream of the evaporator and upstream of the compressor.

TECHNICAL FIELD

The present invention relates to heat pumps, turbines for use with heatpumps and/or generators for use with heat pumps, and in particular, butnot exclusively, to improved refrigeration or air conditioning methodsand apparatus and to turbines and/or generators for use therewith.

BACKGROUND

Present refrigeration cycles reject heat to the atmosphere. In somecases a portion of the energy which would otherwise be rejected may berecovered from the cycle, thereby increasing the overall efficiency.

FIG. 1 shows a diagrammatic representation of a heat pump circuit of theprior art. Hot, high pressure refrigerant liquid enters a throttlingdevice, often referred to as a Tx valve, which reduces its pressure andtemperature at constant enthalpy. The heat absorbing vapour is passedthrough a heat exchanger or “evaporator ” which absorbs heat fromambient temperature air blown across its surfaces by a fan, cooling theair and thereby providing the refrigeration effect and causing it toexpand. The acquisition of heat causes the liquid to flash to vapour andexpand.

The heat laden working fluid vapour is then passed into an accumulatorwhich has an internal structure designed to allow any remaining liquidto boil off prior to entering the compressor.

The energy rich warm working fluid vapour enters a compressor, which asa result of a work input, compresses the vapour thus raising itstemperature and pressure. A significant portion of the work input intothe compressor re-appears as the heat of compression thus superheatingthe working fluid vapour.

The superheated working fluid vapour thus has its temperature elevatedabove that of the ambient temperature of the environment and enters acondenser, which has a structure similar to that of the evaporator. Aheat exchange then occurs between the superheated working fluid vapourand the environment which is at a lower temperature. The heat exchangecontinues until sufficient heat is removed from the working fluid tocause a change of state from hot vapour to hot liquid.

The hot working fluid liquid enters a reservoir, usually referred to asa “receiver” which has a sufficiently large volume to support therequirements of the thermodynamic cycle and withstand the high pressurein the discharge line of the compressor. The hot high pressurerefrigerant liquid then enters the TX valve to complete thethermodynamic cycle.

Air conditioning systems have become a huge draw on electricity power inmany of the major cities of the world and are viewed as an essentialcomponent of many large buildings in order to maintain a level ofenvironmental control within the building. At the same time as airconditioning systems continue to increase in number, it is becomingincreasingly recognised that electricity is a limited resource and insome places demand is exceeding supply or is forecast to in the nearfuture.

It has become important to identify potential areas for saving inelectricity consumption. If any savings can be made in air conditioningsystems, then there is potential to make an overall huge saving in theconsumption of electricity.

The saving of electricity can also lead to savings in power distributioninfrastructure upgrades. Such upgrades are becoming necessary to dealwith increasing peak loads introduced by a rapidly growing airconditioning market.

OBJECT OF THE INVENTION

It is an object of a preferred embodiment of the invention to provideapparatus for a heat pump and/or a heat pump which will increase theutilization of available energy in such apparatus at present.

It is an alternative object of a preferred embodiment of the inventionto provide a method of controlling a heat pump which will increase theefficiency of such apparatus at present.

It is an alternative object of a preferred embodiment of the inventionto provide a method of controlling a turbine and generator which willincrease the efficiency of such apparatus at present.

It is a further alternative object of a preferred embodiment of theinvention to provide a turbine and/or a method of communicating a fluidto a turbine which will increase the utilization of available energyfrom such fluid at present.

It is a still further alternative object to at least provide the publicwith a useful choice.

Other objects of the present invention may become apparent from thefollowing description, which is given by way of example only.

SUMMARY OF THE INVENTION

According to a first aspect of the invention, there is provided athermodynamic cycle including a compressor, a first turbine downstreamof the compressor, a heat exchanger located downstream of the firstturbine and operable to reject heat from the cycle to anotherthermodynamic cycle, an evaporator downstream of the heat exchanger anda second turbine downstream of the evaporator and upstream of thecompressor.

According to a second aspect of the present invention, there is provideda thermodynamic cycle including a compressor, a condenser downstream ofthe compressor, a first turbine downstream of the condenser, anevaporator downstream of the first turbine and a second turbinedownstream of the evaporator and upstream of the compressor.

Preferably, the thermodynamic cycle further includes a heat exchangerlocated between said first turbine and said evaporator, the heatexchanger operable to reject heat to another thermodynamic cycle.

Preferably, at least one of the first turbine and second turbineincludes: a rotor chamber; a rotor rotatable about a central axis withinsaid rotor chamber; at least one nozzle including a nozzle exit forapplying a fluid a fluid supply in the thermodynamic cycle to said rotorto thereby drive said rotor and generate power; at least one exhaustaperture to, in use, exhaust said fluid from said turbine; wherein theflow of said fluid from said at least one nozzle exit is periodicallyinterrupted by at least one flow interrupter means, thereby raising thepressure of said fluid inside said at least one nozzle.

Preferably, the at least one of the first turbine and second turbineincludes at least one fluid storage means between said fluid supply andsaid at least one nozzle.

Preferably, the fluid storage means has a capacity at least equal to adisplacement of the compressor.

Preferably, the at least one flow interrupter means substantially stopsthe flow of said fluid from said at least one nozzle exit until thepressure inside said at least one nozzle rises to a preselected minimumpressure, which is less than or equal to the pressure of the fluidsupply.

Preferably, in use, the flow of said fluid from said at least one nozzleis interrupted by said at least one interrupter means for a periodsufficient to bring said fluid immediately upstream of said at least oneouter nozzle substantially to rest.

Preferably, the rotor has a plurality of channels shaped, positioned anddimensioned to provide a turning moment about said central axis whenrefrigerant from said at least one nozzle enters said channels.

Preferably, the rotor is has a plurality of blades shaped, positionedand dimensioned to provide a turning moment about said central axis whenrefrigerant from said at least one nozzle contacts said blades.

Preferably, the at least one flow interrupter means includes at leastone vane connectable to and moveable with an outer periphery of saidrotor and adapted to interrupt the flow of said fluid out of said atleast one outer nozzle exit when said at least one vane is substantiallyadjacent said at least one nozzle exit.

Preferably, the flow interrupter means includes a plurality of saidvanes substantially evenly spaced apart around said outer periphery ofsaid rotor.

Preferably, the at least one nozzle in use supplies said fluid to saidrotor at a sonic or supersonic velocity.

Preferably, the at least one exhaust aperture includes diffuser andexpander sections to decrease the velocity of said fluid and maintainthe pressure of the fluid flow once it has decelerated to a subsonicvelocity.

Preferably, at least one of the first and second turbines includes arotor including two or more spaced apart rotor windings and a statorincluding a plurality of stator windings about said rotor, wherein atleast two of said stator windings are connected to a controllablecurrent source, each controllable current source operable to energisethe stator windings to which it is connected.

Preferably, each controllable current source is operable to energise thestator windings to which it is connected after the rotor has reached apredetermined velocity.

Preferably, the predetermined velocity is the terminal velocity for thecurrent operating conditions of the turbine.

Preferably, each current source increases or decreases the currentthrough their respective stator windings dependent on a measure of thepower output from the stator windings.

According to another aspect of the present invention, there is provideda method of control for the thermodynamic cycle described in theimmediately preceding four paragraphs including repeatedly measuring thepower output from the stator windings and increasing the current throughthe windings if the current measure of power output is greater than aprevious measure of power output and decreasing the current through thewindings if the current measure of power output is less than a previousmeasure of power output.

According to another aspect of the present invention, there is provideda method of generating power from a thermodynamic cycle including acompressor, a first turbine downstream of the compressor, a heatexchanger located downstream of the first turbine and operable to rejectheat from the cycle to another thermodynamic cycle, an evaporatordownstream of the heat exchanger and a second turbine downstream of theevaporator and upstream of the compressor, wherein the first secondturbines include a rotor and at least one nozzle to apply fluid to therotor to thereby drive said rotor and generate power; the methodincluding providing at least one flow interrupter means to periodicallyinterrupt the flow of said fluid out of said at least one nozzle,thereby raising the pressure of said fluid inside said at least onenozzle to a preselected minimum pressure which is less or equal to saidfluid supply means pressure before resuming the flow of said fluid outof said at least one nozzle.

According to another aspect of the present invention, there is provideda method of generating power from a thermodynamic cycle including acompressor, a condenser downstream of the compressor, a first turbinedownstream of the condenser, an evaporator downstream of the firstturbine and a second turbine downstream of the evaporator and upstreamof the compressor wherein the first second turbines include a rotor andat least one nozzle to apply fluid to the rotor to thereby drive saidrotor and generate power; the method including providing at least oneflow interrupter means to periodically interrupt the flow of said fluidout of said at least one nozzle, thereby raising the pressure of saidfluid inside said at least one nozzle to a preselected minimum pressurewhich is less or equal to said fluid supply means pressure beforeresuming the flow of said fluid out of said at least one nozzle.

Preferably, the preselected minimum pressure is sufficient to cause thefluid to reach the local sonic velocity at a throat of the nozzle.

Preferably, the method includes accelerating fluid exiting said at leastone nozzle to supersonic velocities.

A control system for the thermodynamic cycle described in the precedingparagraphs, the control system including: sensing means for providing ameasure of an output of the thermodynamic cycle; control means for thecompressor, wherein the control means is in communication with saidsensing means to receive as inputs said measure of an output of thethermodynamic cycle and a measure of the work input of the compressor,wherein the control means is operable to compute a measure of efficiencyfrom said inputs and vary the speed of the compressor to maximise saidmeasure of efficiency or to maintain said measure of efficiency at apredetermined level.

Preferably, the control system further includes second control means forthe second turbine and sensing means for providing a measure of thetemperature of a controlled area, wherein the second control meansreceives as a further input said measure of the temperature of acontrolled area, and is operable to open or close the fluid flow paththrough said second turbine in response to sensed variations intemperature in the controlled area in relation to a target measure.

Preferably, the second control means further receives as an input ameasure indicative of the amount of refrigerant in the cycle which isvaporised after an evaporation phase in the cycle and to open or closethe fluid flow path through said second turbine to maintain vaporisedrefrigerant after the evaporation phase.

Preferably, the operation of the second control means to maintainvaporised refrigerant after the evaporation phase is performed after apredetermined delay from the control means opening or closing the fluidflow path through said second turbine in response to said sensedvariations of temperature.

Preferably, the control system includes third control means for acondenser in the thermodynamic cycle, the control system varying theoperation of the condenser to maintain a required level of cooling ofrefrigerant by the condenser.

Preferably, the control means, second control means and third controlmeans is a single microcontroller or microprocessor or a plurality ofmicrocontrollers or microprocessors with at least selectedmicrocontrollers or microprocessors in communication with each other toallow management of the timing of the functions of the control system.

A control system for the thermodynamic cycle described in the precedingparagraphs, the control system including: sensing means for providing ameasure of an output of the thermodynamic cycle; control means for thecompressor, wherein the control means is in communication with saidsensing means to receive as inputs said measure of an output of thethermodynamic cycle and a measure of the work input of the compressor;wherein the control means is operable to compute a measure of efficiencyfrom said inputs and vary the speed of the compressor to maximise saidmeasure of efficiency or to maintain said measure of efficiency at apredetermined level and wherein the control system is operable tocontrol the direct current through the stator windings of said turbine.

Preferably, the control system is operable control the direct currentthrough the stator windings to dynamically maintain the balance of saidturbine when loaded.

Further aspects of the present invention, which should be considered inall its novel aspects, will become apparent from the followingdescription, given by way of example only and with reference to theaccompanying drawings.

BRIEF DESCRIPTION OF DRAWINGS

FIG. 1: Shows a prior art thermodynamic cycle.

FIG. 2: Shows a first thermodynamic cycle according to an aspect of thepresent invention.

FIG. 3: Shows a second thermodynamic cycle according to an aspect of thepresent invention.

FIG. 4: Shows a cross-sectional view of a first turbine according to anaspect of the present invention.

FIG. 5: Shows a cross-sectional view of a second turbine according to anaspect of the present invention.

FIG. 6: Shows an enlarged view of a channel of the turbine of FIG. 5.

FIG. 7: Shows a third thermodynamic cycle illustrating a control systemaccording to an aspect of the present invention.

FIGS. 8-10, 12: Show flow charts of a method of controlling athermodynamic cycle according to aspects of the present invention.

FIG. 11: Shows a diagram of a generator according to an aspect of thepresent invention.

FIG. 13: Shows a flow chart of an initialisation subroutine for thecontrol system.

FIG. 14: Shows a flow chart of a scheduling subroutine for the controlsystem.

BRIEF DESCRIPTION OF PREFERRED EMBODIMENTS OF THE INVENTION

The present invention is described herein with reference to itsapplication to a refrigeration cycle. Those skilled in the art willrecognise that the heat pumping circuit described may have a variety ofuses, for example air conditioning, refrigeration or heating. Thoseskilled in art will also recognise that the term “refrigerant” is usedto describe any working fluid suitable for use in such a circuit orcycle.

A simple refrigeration circuit of the prior art shown in FIG. 1 mayinclude, in order, a compressor, a condenser, a receiver, a throttlingvalve (TX valve), an evaporator and an accumulator. Some embodiments ofthe prior art may combine two of the elements shown in FIG. 1 into asingle device, for example some compressors may also include anaccumulator, but the function of each element is usually present in thecircuit.

The term “turbine” is used herein to describe a device which convertsenergy from a fluid stream into kinetic and/or electrical energy. Thoseskilled in the art will appreciate that where the energy is required inelectrical form the turbine may include a suitable electric powergenerator or alternator.

Referring next to FIG. 2 a heat pump apparatus of the present inventionincludes a first refrigerant circuit 10, which includes in order, afirst compressor 1 a condenser 8, a receiver 2, a TX valve, anevaporator 5 and a turbine 21. The turbine 21 converts energy from therefrigerant into kinetic and/or electrical energy, thereby lowering thetemperature and pressure of the first refrigerant. If required to resultin a suitable density and pressure refrigerant for the turbine, anexpander (not shown) may be provided on one or both of the upstream anddownstream sides of the turbine 21.

In some embodiments the turbine 21 may be designed to avoid cooling therefrigerant to the point where drops of liquid refrigerant form withinthe turbine 21, as this may damage the working surfaces within theturbine 21. In alternative embodiments the turbine 21 may be adapted,for example through the use of appropriately robust materials toconstruct the rotor blades, to allow condensation of the refrigerantwithout damage to the turbine 21.

Those skilled in the art will appreciate which qualities of therefrigerant passing through the first evaporator 5 will affect the heatflow into the first evaporator 5. The refrigerant leaving the firstevaporator 5 passes through a first accumulator 6 before returning tothe first compressor 1. Those skilled in the art will appreciate thatthe receiver 2 and accumulator 6 provides the refrigerant reservoirs forthe circuit. The accumulator 6 is shown in outline to represent theoption that it forms a part of the compressor 1.

Referring to FIG. 3, an alternative heat pump according to the presentinvention is shown, which includes a first refrigerant circuit 300 andsecond refrigerant circuit 400. In a preferred embodiment the secondrefrigerant cycle 400 may include an evaporator 405, accumulator,compressor, condenser, receiver and TX valve (not shown), arranged inthe same order and performing the substantially same function as arefrigeration circuit of the prior art. The second refrigerant may havea boiling point of less than 10° C., more preferably around 0° C. Asuitable second refrigerant may be R22, R134A or R123, although thoseskilled in the art will appreciate that other refrigerants with suitablylow boiling points may be used.

The second refrigerant circuit 400 may be controlled by a control systemas described below with reference to FIG. 7. If required, bothrefrigerant circuits may be controlled by a single controller.

In a preferred embodiment the temperature of the refrigerant enteringthe condenser of the refrigerant circuit 400 may be above 30° C., andpreferably around 60° C. The temperature of the refrigerant entering theevaporator of the refrigerant circuit 400 may be at least 10° C. lowerthan the temperature of the refrigerant entering the condenser 304.

In some embodiments one or more thermoelectric generators positionedbetween a compressor and condenser may be provided in order to generateelectricity. Thermoelectric generators may be particularly useful if therefrigerant used is R123, as the condensing temperature may be as highas 180° C. and the evaporation temperature between 35° C. and 10° C.,thereby providing a large temperature differential.

The cycle 300 includes in clockwise order a compressor 301, condenser307, first expander 302 a, first turbine 302, second expander 302 b, aheat exchanger 304, an evaporator 305 and a second turbine 306.

The expanders may be included on both the input and output sides of theturbine 302 to reduce the density of the working fluid entering theturbine 302, and to assist in maintaining a low pressure at the outputof the turbine 302 after the working fluid returns to a subsonicvelocity. In a preferred embodiment the expander may ensure that thereis no increase in the pressure of the fluid once it has decelerated to asubsonic velocity. Without an expander the pressure at the turbineoutput would otherwise rise and impair the turbine performance.

Expanders (not shown) may also be included on one or both of the inputand output of the second turbine 306. The expanders will include adiffuser if the refrigerant is circulating at supersonic speeds out ofthe turbine 306. The expanders on the inputs of the turbines 302, 306are necessary to lower the density of the working fluid prior toentering the throat of the turbine nozzle. The lower density will allowa larger throat size at the sonic point of the working fluid and hencemaintain a critical minimum mass flow rate so as to avoid any reductionin air conditioning efficiency. Ideally the mass flow rate should be thesame as would be experienced without the introduction of each turbineinto the thermodynamic cycle. The volumetric expansion before the nozzletherefore lowers the density of the working fluid and allows a largerdiameter nozzle throat to be used without impairing either thesubsonic/supersonic transition of the working fluid at the throat or itsmass flow rate.

In two further alternative cycles, one of either the refrigerant cycle400 and condenser 304 may be omitted.

FIG. 4 shows a turbine 21, suitable for use with the heat pump apparatusdescribed in relation to FIGS. 1, 2, 3. The turbine 21 may also be usedin a refrigerant circuit of the prior art, such as the circuit shown inFIG. 1 or in other refrigerant circuit, preferably either immediatelyupstream or downstream of the compressor, with expanders provided aboutthe turbine 21 if necessary. The turbine 21 includes at least one outernozzle 22 mounted in the housing (not shown) of the turbine 21, whichhas a converging/diverging section adapted to accelerate the refrigerantflowing through it to sonic or supersonic speeds.

The turbine 21 is described below with reference to its use as part of aheat pump circuit, such as those described above, in which the workingfluid is refrigerant. The turbine 21 may perform the function of a TXvalve in addition to generating power, allowing a TX valve to be omittedfrom the circuit. Those skilled in the art will appreciate that otherapplications for the turbine 21 are possible and that the working fluidmay in these embodiments be some other suitable gaseous fluid.

The flow from each outer nozzle 22 is periodically interrupted by aninterruption means. Two preferred interruption means are explainedbelow. Those skilled in the relevant arts may be able to identifyalternative means for interrupting the flow from an outer nozzle 22.

A first interruption means may include one or more vanes 7 locatedproximate the outer periphery of the turbine rotor 23 and adapted tosubstantially prevent refrigerant from flowing from an outer nozzle 22when the vane 7 is proximate the outer nozzle outlet 12. Those skilledin the relevant arts will appreciate that the gap between the exit ofthe outer nozzle 22 and the vanes 7 is exaggerated in FIG. 4 and thatthe actual gap will be small enough to interrupt or significantlyinhibit flow from the nozzle 22 when the vanes 7 are adjacent the nozzleexit 12.

A second interruption means 11 may include an electronically operatedvalve proximate the outer nozzle outlet 12. The second interruptionmeans 11 may have an extremely fast response and may, for example, besimilar in operation to an electronically operated common rail dieselinjector.

A refrigerant storage vessel 13 may be located proximate the outernozzle entrance 14. If the compressor supplying refrigerant to the outernozzle 22 is a positive displacement compressor, then the refrigerantstorage vessel 13 may have an internal volume at least equal to a singledisplacement of the first compressor. The refrigerant storage vessel 13may have any capacity greater than the displacement of the compressor.The refrigerant storage vessel 13 may preferably be an insulatedspherical container located as close as possible to the outer nozzleentrance 14.

The vanes 7 and second interruption means 11 may stop the flow ofrefrigerant sufficiently rapidly to cause an adiabatic pressure rise inthe outer nozzle 22 without a corresponding increase in enthalpy. Theflow of refrigerant may be interrupted for a period which issufficiently long for the pressure inside the outer nozzle 22, and morepreferably inside the refrigerant storage vessel 13, to reach apreselected minimum pressure which is less than the pressure supplied bythe first compressor. This pressure may be selected to ensure that whenthe vanes 7 and second interruption means 11 are both open, therefrigerant exits the outer nozzle 22 at sonic or supersonic speeds.

The period of time that each vane 7 stops the flow from the outer nozzle22 depends on the circumference of the turbine rotor 23, the rotationalspeed of the rotor 23 and the length of the vane 7 in thecircumferential direction. In some embodiments this period of time maybe sufficiently long that a second interrupter means 11 is not required.

In other embodiments the second interruption means 11 may be capable ofclosing sufficiently rapidly that the vanes 7 are not necessary, but inmany cases the vanes 7 may provide a relatively simple interruptionmeans, which is capable of closing the outer nozzle outlet 12 at highspeed.

The refrigerant storage vessel 13, vanes 7 and second interruption means11 may assist in increasing the amount of energy recovered from therefrigerant while still allowing sufficient refrigerant to flow toprovide an adequate overall heat absorption effect from a refrigerantcircuit. This may facilitate or assist the omission of a receiver and TXvalve from the refrigeration circuit.

The Applicant believes that when the interruption means closes, the massflow of the working fluid, in this case refrigerant, between the outernozzle 22 and the high pressure source feeding the outer nozzle 22,which in most cases may be a first compressor, may decrease towardszero, and the pressure in the refrigerant storage vessel 13 and outernozzle entrance 14 may rise towards the maximum pressure of thedischarge line of the first compressor. This upward pressure excursionis a function of the decrease in mass flow rate of the fluid. When themass flow rate is zero then the pressure difference across the outernozzle 22 may be substantially zero, therefore the pressure at the outernozzle entrance 14 is at a maximum and the kinetic energy change in therefrigerant is zero and the enthalpy change is zero. Thus, when therefrigerant is stopped the pressure rises at the outer nozzle entrance14 to the maximum value provided by the compressor and the enthalpychange is zero. The Applicant also believes that if the period of timewhen the refrigerant is interrupted is short in comparison to the timein which the refrigerant is allowed to flow, then the deterioration inoverall mass flow in a refrigerant circuit of which the turbine 21 is acomponent will be minimal.

The Applicant further believes that an advantage of stopping the massflow through the outer nozzle 22 is that, if the period of the flowinterruption is sufficiently short and the increase in pressure of therefrigerant occurs substantially adiabatically, there will be no changein the enthalpy of the stationary refrigerant in the outer nozzle 22.Also, if the increase in internal energy during the time when therefrigerant is stationary and the refrigerant is compressed compensatesfor the expansion of the refrigerant and its depletion of work duringthe time when the mass flow is flowing, which may be achieved byproperly selecting the ratio of time during which the refrigerant flowsto time in which the refrigerant is interrupted, then the enthalpyextraction process may become substantially continuous. The Applicantbelieves that this may result in an increased extraction of enthalpyfrom the working fluid over systems of the prior art.

Those skilled in the art will also appreciate that the timing of thesecond interruption means 11 may be controlled by a processing means(not shown). The processing means may receive information on the angularposition of the turbine rotor 23 from any suitable means, but preferablyfrom a hall effect sensor or similar mounted on the turbine housing (notshown), which may sense a suitable index mark on the rotor 23. Theprocessing means may also vary the speed of the turbine rotor 23 byvarying the opening times of the second interrupter 11.

While the turbine rotor 23 is shown having an impulse type bladeconfiguration, the Applicant has found that interrupters as describedabove are also particularly suited to use with other radial type turbinedesigns, for example those used in automotive turbochargers, as is shownin FIG. 11.

Referring now to FIG. 5, an alternative turbine rotor 23A is shown ashaving a plurality of substantially spiral shaped channels 602 leadingto a central exhaust aperture 603. The central exhaust aperture 603 maybe central of the rotor 23A and may extend substantially in thedirection of the central axis of the rotor 23A. The cross-sectional areaof each channel 602 may continuously decrease between an inlet 604 andan outlet 605.

Preferably the ratio of the area of the inlet 604 to the outlet 605 maybe substantially 6:1 in order to promote hypersonic operation with theminimum restriction to the flow of the working fluid.

Referring next to FIG. 6, the centreline 606 of each channel 602 mayintersect a radius 607 of the rotor 23A on at least two points, 608, 609between the inlet 604 and the outlet 605.

A fluid flow, represented by arrows F, may enter a channel 602 throughan inlet 604. As the direction of the fluid F is changed within thechannel 602 the change in momentum of the fluid F may result in aturning force on the rotor 23A. Preferably the turning force may betransmitted to either a suitable electrical energy generator or anyother suitable mechanism which may be powered by a rotating shaft. It ispreferred that the fluid F execute as close as possible to a 180° changein direction within the channel 602 in order to maximise the change inmomentum and therefore the energy imparted to the rotor 23A.

The rotor 23A may be used with an electronic second interrupter lo meansas described above, although those skilled in the art will recognizethat the in some embodiments the spacing 610 between the channelentrances 604 may act as an interrupter means.

FIG. 7 shows an air conditioning/refrigeration cycle, generallyreferenced by arrow 100, according to another aspect of the presentinvention.

Like the cycle 300 shown in FIG. 3, the cycle 100 may differ from airconditioning or refrigeration cycles of the prior art in that the TXvalve and receiver common to the cycles of the prior art, may beomitted. The TX valve is replaced by a turbine 114, which in thisembodiment is located between the condenser 105 and evaporator 122. Anoptional thermoelectric generator 103 may precede the condenser 105.

A second turbine 130 is placed between the output of evaporator 122 andthe accumulator 128. Expander 130 a and 130 b if present are placedabout turbine 130. This is to ensure that the density of the workingfluid entering turbine 130 is sufficiently low, so as to allow asufficiently large diameter nozzle to be used within turbine 130,without impairing the supersonic operation of 130, the mass flow rate ofthe system or its cooling efficiency.

A secondary heat pump cycle referenced by arrow 200 contains a heatexchanger 201 which follows expander 114 c and allows heat to be removedfrom the primary cycle 100, to ensure that the temperature and pressureof the working fluid entering evaporator 122 is sufficiently low toallow the efficient operation of evaporator 122. The secondary cyclecontains all of the essential heat pump components described in theprior art cycle 10 of FIG. 1 with the additional controls referred to inFIG. 7 and described herein for cycle 100.

High pressure working fluid may exit a compressor 101 through acompressor discharge line 102 in a substantially vapour phase and mayenter a thermoelectric generator 103 or may pass straight to a condenser105. The thermoelectric generator 103, if present, may produce a lowvoltage DC output 103 a which may be converted to a high voltage output104 a through a DC to DC converter 104.

The condenser 105 removes heat from the working fluid. The amount ofheat rejected may be controlled by the speed of a condenser fan 106which blows air over the condenser 105. The speed of the condenser fan106 may be determined by a variable speed drive 107, controlled by amaster variable speed drive 109 through a communications link 108. Thevariable speed drive 107 includes suitable software to control the speedof the condenser fan 106.

The master variable speed drive 109 may include thermocouple inputs 110,111 and 112 to provide information on the temperature of refrigerantinto the evaporator (T1), temperature of the refrigerant out of theevaporator (T2) and temperature of air exiting the evaporator (T4)respectively. A further thermocouple (T4 a) and pressure sensor 115 maymeasure the pressure of the temperature and pressure of the workingfluid entering the turbine 114.

By measuring the temperature and pressure of the working fluid enteringthe turbine and selected temperatures in the cycle, the software in themaster variable speed drive 109 may estimate the density of the workingfluid entering the turbine 114 by a software lookup table and may adjustthe speed of the compressor 101 and/or condenser fan 106 and/orevaporator fan 126 to ensure that it is sufficiently low that the vapourpassing through the throat of a ponverging/diverging nozzle 117, whichfeeds the turbine 114, is at a substantially sonic velocity. Expander114 a further reduces the density of the working fluid entering turbine114.

The sonic working fluid exiting the turbine nozzle throat may continueto accelerate in a diverging section of the nozzle 117 until it reachesa supersonic velocity.

The high velocity working fluid drives the turbine rotor. The turbinemay drive a load 121, for example an electric generator, via a suitablecoupling 120.

Acceleration of the working fluid within the nozzle 117, preferably tosonic or supersonic velocities, may cause a fall in its temperature andpressure. Energy may then be removed from the working fluid as a resultof the flowing through the turbine 114.

A mixture of high velocity low pressure working fluid in both vapour andliquid phases is passed into an evaporator 122 via expander 114 c whichis designed to prevent the working fluid pressure from rising as theworking fluid decelerates having had kinetic energy removed from it byturbine 114. If necessary the expander 114 c may also contain a diffuser114 b to cause the velocity of the working fluid to reduce to a subsonicvalue prior to entering expander 114 c.

The evaporator coil 123 may absorb heat from the warmer air 124 outsidethe evaporator 122. The cooled air 125 may be removed from theevaporator 122 by an evaporator fan 126. The speed of the evaporator fan126 may be varied by a further variable speed drive 130 connected to thepower input of the evaporator fan 126 and controlled by the mastervariable speed drive 109 through a communications link 108 a. The speedof the evaporator fan 126 may be varied in response to the drop intemperature of the air 124 flowing over the evaporator 122.

The accumulator 128 may ensure that any remaining liquid phase fluid isevaporated prior to entering the compressor input 129. The accumulator128 may also act as a working fluid reservoir to replace the receiverused by some air conditioning/refrigeration cycles of the prior art.

The master variable speed drive 109 may control the speed of thecompressor 101 to optimise its coefficient of performance (COP),substantially as described herein below, although the TX valve controlwill be omitted due to the elimination of the TX valve from the cycle100.

If the turbine 114 is driving an electrical generator 121 then theelectrical generator 121 may be either of the DC or AC type. Preferablythe generator 121 may be a high voltage DC generator of the order of 670volts output. In the preferred case the DC power output 14B may becoupled into the DC bus bar 109B of the master variable speed drive 109through a diode and capacitor isolation circuit, which may only allowpower to flow in one direction, thus avoiding any feedback of mainspower 150 to the generator 121.

Those skilled in the art will recognise that the air conditioning cyclesdescribed above may be more energy efficient than those of the priorart, due to energy recovered by the turbine and, where used, thethermoelectric generator, as well as the control of the compressor speedto optimize the overall Coefficient of Performance.

FIGS. 8 to 10 show a series of flow diagrams illustrating an example ofthe computational process of the present invention that may be performedto control an air conditioning cycle, such as the cycles describedherein in relation to FIGS. 1, 2, 3, 7, 8 or other cycles includingthose of the prior art if required. The process may be controlled by anysuitable microcontroller, microprocessor or similar having a controloutput to control the drive signal of a motor controller for acompressor. For clarity, in the following description it is assumed thata microcontroller has been used.

Referring to FIG. 8, on power up or before execution of the controlalgorithms, an initialisation routine may be performed in which selectedflags, registers and counters may be initialised, typically by settingto zero if this is required for the particular implementation of thecontrol algorithms.

Referring to FIG. 13, a flow chart illustrating a possibleinitialisation subrouting is shown. The time intervals at which externaldevices (for example the compressor, TX valve, condenser, generatorexcitation) are serviced/optimised are entered as DEL1 to DELn. For theparticular heat pump that is being controlled, a look-up table isdetermined and the entries for target coefficients of performance (COP3to COPn) for the heat pump when operated at a specific temperaturedifferentials across the evaporator ((T1-T3)(1) to (TI-T3)(n)) areentered.

The microprocessor may read the state of a switch SW1. The switch SW1dictates whether the microcontroller automatically schedulesservicing/optimisation of the control parameters for the heat pump. Thecurrent is state of any required flags, counters and registers may alsobe read and then initialised.

A look-up table is then formed from the entered temperaturedifferentials (T1-T3)(1) to (T1-T3)(n) and their associated targetcoefficients of performance COP3 to COPn for use in theservicing/optimisation of the heat pump (see herein below). Finally, themicrocontroller sets a flag that dictates manual or automatic operationbased on the status of the switch SW1.

The microcontroller receives as inputs the temperature of therefrigerant flowing into the evaporator T1, the temperature of therefrigerant leaving the evaporator T2 and the compressor motor powerKW1. The set point for the heat load T3, the required motor speedincrement K2 and required motor speed decrement K3 for the compressorand an air conditioning refrigerant constant K1 are also entered. K1 maybe determined experimentally for the particular air conditioning cycleand represents the increment of heat lifted per degree temperaturechange between T1 and T2.

Having received these inputs, the microcontroller then computes thedifference between T1 and T3. This difference is then used to look up acorresponding coefficient of performance for the heat pump in the storedlook-up table, where the coefficient of performance represents the heatlifted per unit work input.

In an alternative embodiment, instead of working to a target COP, themicrocontroller may increase/decrease the compressor speed to maximisethe COP if the COP for the cycle does not just continually increase withcompressor speed. Those skilled in the relevant arts will alsoappreciate that variables other than the temperature difference acrossthe evaporator may be used if required.

If T1-T3 is less than or equal to zero, the heat pump is not operatingand nothing further is done by the microcontroller, which returns to thestart of the algorithm. If T1-T3 is greater than zero, the actualcoefficient of performance COP2, which is based on the measuredvariables T1, T2 and KW1 is computed according to equation 1:COP 2=K 1|T 1-T 2|/KW 1   equation 1

Other measures relating the output of the cycle to the compressor workinput may be used if required. As herein described, the presentlycontemplated preferred embodiment uses measures of temperaturedifference to provide a measure of the useful heat transferred by thesystem, as temperature measurements may be relatively easily obtained.However, alternative measures of system performance may be used thatrelate the system output to the compressor input.

The computed co-efficient of performance COP2 is then compared to thetarget coefficient of performance COP1. If the value of COP1 is lessthan COP2, the compressor speed is increased by K2. Conversely, if thetarget COP1 is greater than the computed COP2, the motor speed isdecreased by K3. A delay subroutine (not illustrated) is then executedto allow for any lag in the response of the cycle to the change incompressor speed. The required time delay can be determinedexperimentally by forcing adjustments of the compressor speed byincrements of K2 and K3 and measuring the maximum time for the airconditioning cycle to return to steady state conditions. Any suitabledelay subroutine may be used to achieve this delay. The delay subroutineis completed after any control variable is changed before analysing andvarying another control variable to ensure that the system remainsstable and/or to ensure that steady state conditions are used to providemeasures of the inputs to the control algorithms. The execution ofcontrol algorithms may be performed periodically at predetermined timeintervals, continuously with the appropriate time delay between eachcontrol cycle or on a scheduled basis.

FIG. 9 shows diagrammatically a control algorithm to control theoperation of a TX valve, if one is provided in the heat pump. Thecontrol algorithm may also be applied to any controllable device thatperforms the same or similar function to a TX valve.

The microcontroller receives as temperature inputs the unsaturatedtemperature of the air exiting the evaporator T4 and a constant T5representing a superheat temperature value added to the temperature ofthe working fluid at the evaporator output. It also receives a pressureinput P1 representing the pressure of the working fluid at theevaporator output, a measurement of the current status of a TX valve orequivalent TX1, and set steps K4 and K5 for incrementing anddecrementing the operation of the TX valve respectively.

The microcontroller computes T6 as the sum of T4 and T5 and computes T7as the product of P1 with a constant K6, which facilitates theconversion of pressure to temperature of the working fluid. If thetemperature T6 is less than T7, the TX valve is opened by increment K4and if the temperature T6 is greater than T7, the TX valve is closed byincrement K5. Otherwise, the TX valve is maintained in its currentposition. The incremental and decremental step size may optionally bethe same (K4=K5). A delay subroutine is then executed in order to allowthe cycle to reach a steady state or near steady state before anyfurther action is taken.

With variation of the TX valve setting, it may be advantageous to checkthat the TX valve is still operating so that the refrigerant in thesuction line of the compressor after the evaporator is sufficientlysuper heated to be at the vapour state. Therefore, each time when thedelay subroutine following variation of the TX valve is invoked, themicrocontroller may perform an additional check on the operation of theTX valve. This check may only be necessary if the control over thelimits of operation of the TX valve is not already present as part ofthe TX valve and if the existing control algorithms do not bound the TXvalve within an acceptable operating range.

With variations in the compressor speed and TX valve opening, theoperation of the condenser will also vary. Therefore, the controller mayalso control the drive fan to a condenser. This process is shown in FIG.10.

The temperature inputs to the algorithm are T1 and T3 as defined hereinabove, the liquid line temperature T8, measured at a predetermined pointin the heat pump, typically at a point immediately following thecondenser and the target temperature for the liquid line temperatureT10. The step size for an increment in condenser fan speed K7 and stepsize for an increment in condenser fan speed K8 are also inputs to thealgorithm together with the current condenser fan speed CFS1, minimumcondenser fan speed CFSmin and maximum condenser fan speed CFSmax.Although the steps that use CFSmin and CFSmax are not illustrated inFIG. 11 the values of CFSmin and CFSmax bound the allowable speed of thecompressor fan.

The microcontroller first calculates T11 as the difference between T3and T1 and terminates the control algorithm for the condenser fan speedif T3 is greater than or equal to T1. If T3 is less than T1, the cycleis operational and heat extracted by the condenser. The microcontrollerthen calculates T12 as the difference of T10 and T8 and if the targettemperature T10 is less than the actual temperature T8 the currentcompressor speed CFSI is increased by K7 and if T10 is greater than T8the current compressor speed is decreased by K8. A further time delay isinvoked after variation of the condenser fan operation.

The microprocessor may also vary the timing of the second interrupter 11to optimize a selected parameter of each refrigerant circuit. In someembodiments the heat absorbed by an evaporator may be the selectedparameter, while in other embodiments the total power input one or moreof the compressors may be the selected parameter.

FIG. 14 shows diagrammatically a control algorithm for the scheduling ofcontrol/optimisation algorithms described herein above. A table of timeparameters is stored in memory, which specifies when each algorithm isto be executed. This table of time parameters will be entered by theheat pump administrator. On power up, a pointer is set to an initialvalue in the table of time parameters and the clock started. The tableof time parameters lists sequentially all of the control algorithms, atime delay variable that indicates the time delay that should occurbetween each execution for that control algorithm and an addressindicating where the control algorithm can be found in memory.

The microcontroller reads the current time of the real time clock andadds the time delay indicated in the time parameters table to give itthe current servicing time. The current servicing time is then read andcompared with the real time clock. The process continually cycles arounda loop, checking the real time against the current servicing time foreach algorithm, until the real time clock reaches the current servicingtime for an algorithm. When this occurs, the microprocessor exits theloop, reads the start address for the algorithm from the time parameterstable and executes the algorithm. After the algorithm has been executed,the microprocessor returns to the loop as indicated by “return” in FIG.14.

The rotors in the generators of the heat pump may operate at highrotational speeds. For example the generators and heat pump may bedesigned so that the rotors revolve at 15000 rpm or more. To maintainthe performance of the generator at high revolution speeds, it isnecessary to balance the rotating group (turbine, rotor, shaft andbearing system). Also, sealing the rotor and generator into therefrigerant cycle may avoid problems with losses and reliability oftransferring power of the cycle through a shaft. Furthermore, if a fixedmagnet rotor is used, sensitive balancing becomes difficult due to themagnetic field about the rotor and the ferromagnetic components of therig become magnetised and if a sudden load is applied to the generator,the resulting force can unbalance the rotor.

The generator of the present invention includes a rotor that isnon-magnetic and can not become magnetised. The rotor may, for example,be produced from Lycore 150 electrical sheet steel. The electric fieldemanating from the rotor is controlled by coils provided on the rotorwound on high permeability F5 ferrite rod formers. Other suitablematerials may be used.

The turbine components in close proximity to the rotor and the casingfor the rotor may both be constructed from a suitable plastic resistantto the high stresses applied in the generator. These componentstherefore do not interfere with the electrical field from the rotor orthe electrical field from energised stator windings. The stator windingsare wound onto a toroidal core about the plastic casing. The toroidalcore may be Lycore 150 electrical sheet steel or more preferably a highpermeability specially moulded ferrite former of F5 ferrite orequivalent.

FIGS. 11A-D show a turbine generator generally referenced by arrow 500.The entire generator 500 may be sealed within the air conditioningcycle, FIG. 11A shows a top view of a turbine generator 500, with coversremoved for clarity and FIG. 11B shows a section though line BB in FIG.11A. The turbine generator 500 includes a turbine housing 501, a statorsupport housing 502 supporting a stator 504 and cover plates 503A-D.FIGS. 11C and 11D show a section through lines CC and DD in FIG. 11Brespectively. The turbine housing 501 contains a turbine 505 including arotor 506 and a nozzle 507 held in place by a nozzle retainer 508. Thenozzle 507 is supplied with refrigerant through an inlet pipe 509. Thegenerator rotor 510 includes four rotor coils 511-514 forming afour-pole rotor 510. The coils 511-514 may have their ends shortedtogether or connected by a resistive element which impedance/resistanceincreases with temperature to provide current limiting to protect thewindings of the rotor. The coils may, for example, be formed from 1 mmcopper and have 135 turns about a 19 mm F5 ferrite former. However, aswill be appreciated by those skilled in the relevant arts, the number ofwindings in both the generator rotor 510 and stator 504, the core usedfor the windings, the air gap between the generator rotor 510 and statorwindings and the number of poles provided on the generator rotor 510 canbe varied according to the requirements for the generator 500. Theturbine rotor 506 preferably has interrupters as described above withreference to FIG. 4, and may have a blade structure as described hereinin relation to FIG. 4 or 5.

The windings of the stator 504 may be wired together in adjacent groupsof two or more windings. The AC outputs of each winding group areconnected to other groups at 90 degree intervals for the four pole rotor510. The winding groups are each connected to a controlled DC generator(not shown) that is operable to feed a constant direct current thoughthe stator windings. Capacitors isolate the windings and DC generatorfrom the AC output. Winding groups are energised with a direct currentcreating alternate north and south pole pairs about the rotor, which maybe at 90 degree intervals, with the like fields being placed oppositeeach other at 180 degree intervals. The electric field is thereforebalanced around the rotor 510 and can if necessary be adjusted tocorrect any imbalance in the rotor 510 in response to any imbalance thatmay be detected during operation. The other stator windings will nothave a DC generator connected to them. By way of example, there may be atotal of 18 coil groups in the stator, with four connected to DCgenerators. Two, three or more than four stator windings connected to DCgenerators may be provided if required.

The polarity of the DC current can be reversed periodically to ensurethat the ferromagnetic components in the turbine 500 do not acquire apermanent magnetic bias.

Turbines of the prior art have operating speed and torquecharacteristics that are fixed and can not be controlled without loss ofperformance. However, the turbine 500 of the present invention allowsdynamic control of the strength of the exciting field, changing thecharacteristics of the generator so that the turbine 500 can be operatedat the most favourable speed and torque to maintain operation withinfixed parameters. For application to the turbines in the heat pumpsdescribed herein, the turbine 500 of the present invention may be usedto maintain supersonic operation.

When the turbine 500 reaches its terminal velocity, the DC currentgenerators are activated, causing an electric field to be generated bythe stator windings connected to the generator, which generates an ACcurrent in the coils of the rotor 510 as the rotor 510 rotates. ACcurrent is then generated in the stator windings, which are fed to thegenerator output. The AC output may be rectified and if the generatorforms part of a heat pump, the energy may be used to partially power acompressor in the heat pump.

FIG. 12 shows diagrammatically a control algorithm for the statorwindings. The control algorithm shown in FIG. 12 is used after the rotor510 has been brought up to speed and direct current is being fed throughthe stator windings. The total current output IT and total voltageoutput VT from the stator is measured. This may be achieved by takingmeasurements of the current output 11-In and voltage output V1-Vn foreach stator winding group. The total power output is computed as theproduct of IT and VT. This is compared to the previous power output. Ifthe previous power output was less than the current power output, thedirect current through the stator windings is increased by apredetermined step size. If the previous power output was more than thecurrent power output, the direct current through the stator windings isdecreased by a predetermined step size. Those skilled in the art willappreciate that the algorithm illustrated in FIG. 12 may use to controlmultiple target generators.

Where in the foregoing description, reference has been made to specificcomponents or integers of the invention having known equivalents thensuch equivalents are herein incorporated as if individually set forth.

Although this invention has been described by way of example and withreference to possible embodiments thereof, it is to be understood thatmodifications or improvements may be made thereto without departing fromthe scope of the invention as defined in the appended claims.

1. A thermodynamic cycle including a compressor, a first turbinedownstream of the compressor, a heat exchanger located downstream of thefirst turbine and operable to reject heat from the cycle to anotherthermodynamic cycle, an evaporator downstream of the heat exchanger anda second turbine downstream of the evaporator and upstream of thecompressor.
 2. A thermodynamic cycle including a compressor, a condenserdownstream of the compressor, a first turbine downstream of thecondenser, an evaporator downstream of the first turbine and a secondturbine downstream of the evaporator and upstream of the compressor. 3.The thermodynamic cycle of claim 2 further including a heat exchangerlocated between said first turbine and said evaporator, the heatexchanger operable to reject heat to another thermodynamic cycle.
 4. Thethermodynamic cycle of any one of claims 1 to 3, wherein at least one ofthe first turbine and second turbine includes: a rotor chamber; a rotorrotatable about a central axis within said rotor chamber; at least onenozzle including a nozzle exit for applying a fluid a fluid supply inthe thermodynamic cycle to said rotor to thereby drive said rotor andgenerate power; at least one exhaust aperture to, in use, exhaust saidfluid from said turbine; wherein the flow of said fluid from said atleast one nozzle exit is periodically interrupted by at least one flowinterrupter means, thereby raising the pressure of said fluid insidesaid at least one nozzle.
 5. The thermodynamic cycle of claim 4, whereinthe at least one of the first turbine and second turbine includes atleast one fluid storage means between said fluid supply and said atleast one nozzle.
 6. The thermodynamic cycle of claim 5, wherein saidfluid storage means has a capacity at least equal to a displacement ofthe compressor.
 7. The thermodynamic cycle of any one of claims 4 to 6,wherein said at least one flow interrupter means substantially stops theflow of said fluid from said at least one nozzle exit until the pressureinside said at least one nozzle rises to a preselected minimum pressure,which is less than or equal to the pressure of the fluid supply.
 8. Thethermodynamic cycle of any one of claims 4 to 7, wherein in use, saidflow of said fluid from said at least one nozzle is interrupted by saidat least one interrupted means for a period sufficient to bring saidfluid immediately upstream of said at least one outer nozzlesubstantially to rest.
 9. The thermodynamic cycle of any one of claims 4to 8, wherein said rotor has a plurality of channels shaped, positionedand dimensioned to provide a turning moment about said central axis whenrefrigerant from said at least one nozzle enters'said channels.
 10. Thethermodynamic cycle of any one of claims 4 to 9, wherein said rotor ishas a plurality of blades shaped, positioned and dimensioned to providea turning moment about said central axis when refrigerant from said atleast one nozzle contacts said blades.
 11. The thermodynamic cycle ofany one of claims 4 to 10, wherein said at least one flow interruptermeans includes at least one vane connectable to and moveable with anouter periphery of said rotor and adapted to interrupt the flow of saidfluid out of said at least one outer nozzle exit when said at least onevane is substantially adjacent said at least one nozzle exit.
 12. Thethermodynamic cycle of claim 11, wherein said flow interrupter meansincludes a plurality of said vanes substantially evenly spaced apartaround said outer periphery of said rotor.
 13. The turbine of any one ofclaims 4 to 12, wherein said at least one nozzle in use supplies saidfluid to said rotor at a sonic or supersonic velocity.
 14. Thethermodynamic cycle of claim 13, wherein said at least one exhaustaperture includes diffuser and expander sections to decrease thevelocity of said fluid and maintain the pressure of the fluid flow onceit has decelerated to a subsonic velocity.
 15. The thermodynamic cycleof any one of claims 1 to 14, wherein at least one of the first andsecond turbines includes a rotor including two or more spaced apartrotor windings and a stator including a plurality of stator windingsabout said rotor, wherein at least two of said stator windings areconnected to a controllable current source, each controllable currentsource operable to energise the stator windings to which it isconnected.
 16. The thermodynamic cycle of claim 15, wherein eachcontrollable current source is operable to energise the stator windingsto which it is connected after the rotor has reached a predeterminedvelocity.
 17. The thermodynamic cycle of claim 16, wherein thepredetermined velocity is the terminal velocity for the currentoperating conditions of the turbine.
 18. The thermodynamic cycle of anyone of claims 15 to 17, wherein each current source increases ordecreases the current through their respective stator windings dependenton a measure of the power output from the stator windings.
 19. A methodof control for the thermodynamic cycle claimed in any one of claims 15to 18 including repeatedly measuring the power output from the statorwindings and increasing the current through the windings if the currentmeasure of power output is greater than a previous measure of poweroutput and decreasing the current through the windings if the currentmeasure of power output is less than a previous measure of power output.20. A method of generating power from a thermodynamic cycle including acompressor, a first turbine downstream of the compressor, a heatexchanger located downstream of the first turbine and operable to rejectheat from the cycle to another thermodynamic cycle, an evaporatordownstream of the heat exchanger and a second turbine downstream of theevaporator and upstream of the compressor, wherein the first secondturbines include a rotor and at least one nozzle to apply fluid to therotor to thereby drive said rotor and generate power; the methodincluding providing at least one flow interrupter means to periodicallyinterrupt the flow of said fluid out of said at least one nozzle,thereby raising the pressure of said fluid inside said at least onenozzle to a preselected minimum pressure which is less or equal to saidfluid supply means pressure before resuming the flow of said fluid outof said at least one nozzle.
 21. A method of generating power from athermodynamic cycle including a compressor, a condenser downstream ofthe compressor, a first turbine downstream of the condenser, anevaporator downstream of the first turbine and a second turbinedownstream of the evaporator and upstream of the compressor wherein thefirst second turbines include a rotor and at least one nozzle to applyfluid to the rotor to thereby drive said rotor and generate power; themethod including providing at least one flow interrupter means toperiodically interrupt the flow of said fluid out of said at least onenozzle, thereby raising the pressure of said fluid inside said at leastone nozzle to a preselected minimum pressure which is less or equal tosaid fluid supply means pressure before resuming the flow of said fluidout of said at least one nozzle.
 22. The method of claim 20 or claim 21,wherein said preselected minimum pressure is sufficient to cause thefluid to reach the local sonic velocity at a throat of the nozzle. 23.The method of claim 22, including accelerating fluid exiting said atleast one nozzle to supersonic velocities.
 24. A control system for thethermodynamic cycle claimed in any one of claims 1 to 18, the controlsystem including: sensing means for providing a measure of an output ofthe thermodynamic control means for the compressor, wherein the controlmeans is in communication with said sensing means to receive as inputssaid measure of an output of the thermodynamic cycle and a measure ofthe work input of the compressor; wherein the control means is operableto compute a measure of efficiency from said inputs and vary the speedof the compressor to maximise said measure of efficiency or to maintainsaid measure of efficiency at a predetermined level.
 25. The controlsystem of claim 24, further including second control means for thesecond turbine and sensing means for providing a measure of thetemperature of a controlled area, wherein the second control meansreceives as a further input said measure of the temperature of acontrolled area, and is operable to open or close the fluid flow paththrough said second turbine in response to sensed variations intemperature in the controlled area in relation to a target measure. 26.The control system of claim 24 or claim 25, wherein the second controlmeans further receives as an input a measure indicative of the amount ofrefrigerant in the cycle which is vaporised after an evaporation phasein the cycle and to open or close the fluid flow path through saidsecond turbine to maintain vaporised refrigerant after the evaporationphase.
 27. The control system of any one of claims 24 to 26, wherein theoperation of the second control means to maintain vaporised refrigerantafter the evaporation phase is performed after a predetermined delayfrom the control means opening or closing the fluid flow path throughsaid second turbine in response to said sensed variations oftemperature.
 28. The control system of any one of claims 24 to 27including third control means for a condenser in the thermodynamiccycle, the control system varying the operation of the condenser tomaintain a required level of cooling of refrigerant by the condenser.29. The control system of claim 28, wherein the control means, secondcontrol means and third control means is a single microcontroller ormicroprocessor or a plurality of microcontrollers or microprocessorswith at least selected microcontrollers or microprocessors incommunication with each other to allow management of the timing of thefunctions of the control system.
 30. A control system for thethermodynamic cycle claimed in any one of claims 15 to 17, the controlsystem including: sensing means for providing a measure of an output ofthe thermodynamic cycle; control means for the compressor, wherein thecontrol means is in communication with said sensing means to receive asinputs said measure of an output of the thermodynamic cycle and ameasure of the work input of the compressor; wherein the control meansis operable to compute a measure of efficiency from said inputs and varythe speed of the compressor to maximise said measure of efficiency or tomaintain said measure of efficiency at a predetermined level and whereinthe control system is operable to control the direct current through thestator windings of said turbine.
 31. The control system of claim 30,operable control the direct current through the stator windings todynamically maintain the balance of said turbine when loaded.